Rotary blood pump and control system therefor

ABSTRACT

A pump assembly and estimation and control system therefor, the pump adapted for continuous flow pumping of blood. In a particular form, the pump is a centrifugal pump wherein the impeller is entirely sealed within the pump housing and is exclusively hydrodynamically suspended therein against movement in three translational and two rotational degrees of freedom as the impeller rotates within the fluid urged by electromagnetic means external to the pump cavity. Hydrodynamic suspension is assisted by the impeller having deformities therein such as blades with surfaces tapered from the leading edges to the trailing edges of bottom and top surfaces thereof.

RELATED APPLICATIONS

This application is a divisional application of U.S. application Ser.No. 09/980,682 filed Aug. 15, 2002 which is a national phase applicationof the International Application PCT/AU00/00355 filed Apr. 20, 2000 andclaims the priority benefits of U.S. application Ser. No. 09/299,038filed Apr. 23, 1999 (issued Jun. 26, 2001 as U.S. Pat. No. 6,250,880)and Australian PP 9959 filed Apr. 23, 1999, all of which areincorporated herein by reference in their entireties.

BACKGROUND OF THE INVENTION

1. Field of the Invention

This invention relates to rotary pumps adapted, but not exclusively, foruse as artificial hearts or ventricular assist devices and, inparticular, discloses in preferred forms a seal-less shaft-less pumpfeaturing open or closed (shrouded) impeller blades with at least partsof the impeller used as hydrodynamic thrust bearings and withelectromagnetic torque provided by the interaction between magnetsembedded in the blades or shroud and a rotating fixed relative to thecurrent pattern generated in coils pump housing.

In addition, a non-contact estimation and control system is describedfor use in conjunction with the rotary pumps of the invention.

2. Description of the Related Art

This invention relates to the art of continuous or pulsatile flow rotarypumps and, in particular, to electrically driven pumps suitable for usealthough not exclusively as an artificial heart or ventricular assistdevice. For permanent implantation in a human patient, such pumps shouldideally have the following characteristics: no leakage of fluids into orfrom the bloodstream; parts exposed to minimal or no wear; minimumresidence time of blood in pump to avoid thrombosis (clotting); minimumshear stress on blood to avoid blood cell damage such as haemolysis;maximum efficiency to maximise battery duration and minimise bloodheating; and absolute reliability.

Several of these characteristics are very difficult to meet in aconventional pump configuration including a seal, i.e. with an impellermounted on a shaft which penetrates a wall of the pumping cavity, asexemplified by the blood pumps referred to in U.S. Pat. No. 3,957,389 toRafferty et al., U.S. Pat. No. 4,625,712 to Wampler, and U.S. Pat. No.5,275,580 to Yamazaki. Two main disadvantages of such pumps are firstlythat the seal needed on the shaft may leak, especially after wear, andsecondly that the rotor of the motor providing the shaft torque remainsto be supported, with mechanical bearings such as ball-bearingsprecluded due to wear. Some designs, such as U.S. Pat. No. 4,625,712 toWampler and U.S. Pat. No. 4,908,012 to Moise et al., have overcome theseproblems simultaneously by combining the seal and the bearing into onehydrodynamic bearing, but in order to prevent long residence times theyhave had to introduce means to continuously supply a blood-compatiblebearing purge fluid via a percutaneous tube.

In seal-less designs, blood is permitted to flow through the gap in themotor, which is usually of the brushless DC type, i.e. comprising arotor including permanent magnets and a stator in which an electriccurrent pattern is made to rotate synchronously with the rotor. Suchdesigns can be classified according to the means by which the rotor issuspended: contact bearings, magnetic bearings or hydrodynamic bearings,though some designs use two of these means.

Contact or pivot bearings, as exemplified by U.S. Pat. No. 5,527,159 toBozeman et al. and U.S. Pat. No. 5,399,074 to Nose et al., havepotential problems due to wear, and cause very high localised heatingand shearing of the blood, which can cause deposition and denaturationof plasma −3 proteins, with the risk of embolisation and bearingseizure.

Magnetic bearings, as exemplified by U.S. Pat. No. 5,350,283 to Nakazekiet al., U.S. Pat. No. 5,326,344 to Bramm et al. and U.S. Pat. No.4,779,614 to Moise et al., offer contactless suspension, but requirerotor position measurement and active control of electric current forstabilisation of the position in at least one direction, according toEamshaw's theorem. Position measurement and feedback control introducesignificant complexity, increasing the failure risk. Power use by thecontrol current implies reduced overall efficiency. Furthermore, size,mass, component count and cost are all increased.

U.S. Pat. No. 5,507,629 to Jarvik claims to have found a configurationcircumventing Earnshaw's Theorem and thus requiring only passivemagnetic bearings, but this is doubtful and contact axial bearings areincluded in any case. Similarly, passive radial magnetic bearings and apivot point are employed in U.S. Pat. No. 5,443,503 to Yamane.

Prior to the present invention, pumps employing hydrodynamic suspension,such as U.S. Pat. No. 5,211,546 to Isaacson et al. and U.S. Pat. No.5,324,177 to Golding et al., have used journal bearings, in which radialsuspension is provided by the fluid motion between two cylinders inrelative rotation, an inner cylinder lying within and slightly off axisto a slightly larger diameter outer cylinder. Axial suspension isprovided magnetically in U.S. Pat. No. 5,324,177 and by either a contactbearing or a hydrodynamic thrust bearing in U.S. Pat. No. 5,211,546.

U.S. Pat. No. 4,944,748 discloses a magnetically suspended impellerwithin a pump. It does not disclose an exclusively hydrodynamicallysuspended impeller within a pump.

U.S. Pat. No. 4,688,998 again discloses a magnetically suspendedimpeller. It does not disclose a hydrodynamically suspended impeller,much less an exclusively hydrodynamically suspended impeller within apump.

WO 91/19103 to NU-TECH discloses an axial flow blood pump having ahydrodynamically suspended rotor assisted by magnetic or mechanicalstabilisation.

U.S. Pat. No. 5,112,200 to NU-TECH discloses hydrodynamic support in atleast one dimension, but utilising prior art hydrodynamic lift surfaceswhich do not include the deformed surfaces of the present invention.

WO 94/13955 discloses a fluid pump which relies on a magneticallylevitated impeller.

U.S. Pat. No. 4,382,199 to NU-TECH discloses a rotor and impellercombination which employs “squeeze film effects, dash pot effects andhydrodynamic effects, all of which combine and co-operate to preventmetal-to-metal contact between the rotor and the stator and to lubricatethe rotor as it rotates within the stator” (column 6). There is nodisclosure of exclusive hydrodynamic support in all dimensions by theuse of deformed surfaces.

A purging flow is needed through the journal bearing, a high shearregion, in order to remove dissipated heat and to prevent long fluidresidence time. It would be inefficient to pass all the fluid throughthe bearing gap, of small cross-sectional area, as this would demand anexcessive pressure drop across the bearing. Instead a leakage path isgenerally provided from the high pressure pump outlet, through thebearings and back to the low pressure pump inlet, implying a smallreduction in outflow and pumping efficiency. U.S. Pat. No. 5,324,177provides a combination of additional means to increase the purge flow,namely helical grooves in one of the bearing surfaces, and a smalladditional set of impellers.

U.S. Pat. No. 5,211,546 provides 10 embodiments with various locationsof cylindrical bearing surfaces. One of these embodiments, the third,features a single journal bearing and a contact axial bearing.

Embodiments of the present invention offer a relatively low cost and/orrelatively low complexity means of suspending the rotor of a seal-lessblood pump, thereby overcoming or ameliorating the problems of existingdevices mentioned above.

SUMMARY OF THE INVENTION

Accordingly, in one broad form of the invention there is provided arotary blood pump for use in a heart assist device or like device, saidpump having an impeller suspended in use within a pump housingexclusively by hydrodynamic thrust forces generated by relative movementof said impeller with respect to and within said pump housing; andwherein at least one of said impeller of said housing includes at leasta first deformed surface lying on at least part of a first face and asecond deformed surface lying on at least part of a second face which,in use, move relative to respective facing surfaces on the other of saidimpeller or said housing thereby to form at least two relatively movingsurfaces pairs which generate relative hydrodynamic thrust between saidimpeller and said housing which includes everywhere a localized thrustcomponent substantially and everywhere normal to the plane of movementof said first deformed surface and said second deformed surface withrespect to said facing surfaces; and wherein the combined effect of thelocalized normal forces generated on the surfaces of said impeller is toproduce resistive forces against movement in three translational and tworotational degrees of freedom.

In yet a further broad form of the invention there is provided anestimation and control system for a pump; said pump of the type havingan impeller located within a pump cavity in a pump housing; said housinghaving a fluid inlet in fluid communication with said pump cavity; saidimpeller urged to rotate about an impeller axis so as to cause fluid tobe urged from said inlet through said pump cavity to said pump outlet;said impeller urged to rotate by impeller urging means; said impellersupported for rotational movement by impeller support means; saidimpeller maintained at or near a predetermined speed of rotation bycontrol means acting on said impeller urging means; said control meansreceiving as input variables a first input variable comprising powerconsumed by said urging means; said control means receiving a secondinput variable comprising actual speed of rotation of said impeller;said control means thereby estimating head across the pump and/or rateof flow of said fluid to an approximation of predetermined accuracyrelying on signals available from said urging means; said control systemadapted to maintain speed of rotation of said impeller within a rangewhereby said impeller, in use, substantially resists five degrees offreedom of movement with respect to said pump housing predominantlywithout any external intervention from said control system to controlthe position of said impeller with respect to said housing.

In yet a further broad form of the invention there is provided a rotaryblood pump and an estimation and control system therefor, said pumphaving an impeller suspended hydrodynamically within a pump housing bythrust forces generated by the impeller during movement in use of theimpeller as it rotates about an impeller axis; said estimation andcontrol system of the type described above.

In yet a further broad form of the invention there is provided a rotaryblood pump having a housing within which an impeller acts by rotationabout an impeller axis to cause a pressure differential between an inletside of the pump housing of said pump and an outlet side of the pumphousing of said pump; said impeller suspended hydrodynamically by thrustforces generated by the impeller during movement in use of the impeller;said pump controlled by the estimation and control system as describedabove.

In yet a further broad form of the invention there is provided aseal-less, shaft-less pump comprising a housing defining a chambertherein and having a liquid inlet to said chamber and a liquid outletfrom said chamber; said pump further including an impeller locatedwithin said chamber; the arrangement between said impeller, said inlet,said outlet, and the internal walls of said chamber being such that uponrotation of said impeller about an impeller axis relative to saidhousing liquid is urged from said inlet through said chamber to saidoutlet; and wherein thrust forces are generated by relative movement ofsaid impeller with respect to said housing; said pump controlled by theestimation and control system as described above.

In yet a further broad form of the invention there is provided a pumphaving a housing within which an impeller acts by rotation about an axisto cause a pressure differential between an inlet side of a housing ofsaid pump and an outlet side of the housing of said pump; said impellersuspended exclusively hydrodynamically in at least one of a radial oraxial direction by thrust forces generated by the impeller duringmovement in use of the impeller; said pump controlled by the estimationand control system as described above.

In yet a further broad form of the invention there is provided a methodof hydrodynamically suspending and controlling an impeller within arotary pump for support in at lest one of a radial or axial direction;said method comprising incorporating a deformed surface in at least partof said impeller so that, in use, a thrust is created between saiddeformed surface and the adjacent pump casing during relative movementtherebetween; said method further including the step of maintainingspeed of rotation of said impeller within a range whereby said impeller,in use, substantially resists five degrees of freedom of movement withrespect to said pump housing without any external intervention.

In yet a further broad form of the invention there is provided anestimation and control system for a pump; said pump of the type havingan impeller located within a pump cavity in a pump housing; said housinghaving a fluid inlet in fluid communication with said cavity; saidhousing having a fluid outlet in fluid communication with said pumpcavity; said impeller urged to rotate about an impeller axis so as tocause fluid to be urged from said inlet through said pump cavity to saidpump outlet; said impeller urged to rotate by impeller urging means;said impeller supported for rotational movement by impeller supportmeans; said pump maintained at or near a predetermined operating pointby control means acting on said impeller urging means; said controlmeans receiving as input at least a first input variable derived fromsaid urging means; said control means receiving at least a second inputvariable also derived from said urging means; said control means therebycalculating an estimate of said operating point to an approximation ofpredetermined accuracy relying on signals available from said urgingmeans; said control means controlling said pump by comparing saidpredetermined operating point with said estimate of said operatingpoint; and wherein instantaneous pump speed and electrical input powerare allowed to be modulated by the heart, in use, by appropriateselection of a control time constant.

In yet a further broad form of the invention there is provided aphysiological controller for use in association with a pump; saidcontroller monitoring estimated flow of fluid within said pump andpressure across said pump by non-contact means thereby to control speedof rotation of an impeller within said pump; and wherein said controllerpermits impeller speed to vary under a pulsating fluid load thereby toassist in calculation and adjustment of impeller speed set point.

Preferably said pump comprises a ventricular assist device adapted toassist operation of a ventricle of a hear and wherein said control meansadjusts pump output so that, in alternation fashion, said ventricle inconjunction with said aortic valve is allowed to eject blood over apredetermined number of cardiac cycles and then said ventricle inconjunction with said aortic valve is caused to eject blood over afollowing predetermined number of cardiac cycles.

In yet a further broad form of the invention there is provided anestimation and control system for a pump; said pump of the type havingan impeller located within a pump cavity in a pump housing; said housinghaving a fluid inlet in fluid communication with said cavity; saidhousing having a fluid outlet in fluid communication with said pumpcavity; said impeller urged to rotate about an impeller axis so as tocause fluid to be urged from said inlet through said pump cavity to saidpump outlet; said impeller urged to rotate by impeller urging means;said impeller supported for rotational movement by impeller supportmeans; said pump maintained at or near a predetermined operating pointby control means acting on said impeller urging means; said controlmeans receiving as input variables at least a first input variablederived from said urging means; said control means receiving at least asecond input variable also derived from said urging means; said controlmeans thereby calculating an estimate of said operating point to anapproximation of predetermined accuracy relying on signals availablefrom said urging means; said control means controlling said pump bycomparing said predetermined operating point with said estimate of saidoperating point; and wherein said pump is arranged to operate accordingto a relatively flat HQ characteristic.

Preferable there is no inflexion point of said HQ characteristic at ornear said predetermined operating point.

Preferably said pump includes near-radial off-flow from said impeller.

Preferably said pump has a low specific speed.

Preferably said pump is a low specific speed pump.

Preferably said pump is specified in a range of 100-2000 rev/min(gal/min)^(1/2)ft^(−3/4).

Preferably said pump has a specific speed of approximately 900-1000rev/min (gal/min)^(1/2)ft^(−3/4).

Preferably instantaneous pump speed and electrical input power areallowed to be modulated by the heart, in use, by appropriate selectionof time constant.

Preferably the time constant of the control system is greater than therotational, inertial time constant of the impeller.

Preferably said time constant is at least one cardiac cycle.

Preferably said first input variable comprises instantaneous pump speed.

Preferably said second input variable comprises electrical input powerto said impeller urging means.

Preferably said pump is arranged to operate according to a relativelyflat HQ characteristic.

In a particular preferred form said HQ characteristic is sufficientlyflat that head will remain constant to a sufficient approximation over apredetermined operating range whereby, over said operating rangewhereby, over said operating range, said system can assume that pumpspeed will be proportional to flow rate.

Preferably said predetermined operating point is calculated so as tomaintain minimum pump speed such that the minimum head pressure acrossthe pump does not increase.

Preferably said system ensures that minimum pump speed is always greaterthan or equal to the minimum speed at which non-regurgitant flow willoccur.

Preferably the speed at which regurgitant or negative flow will begin tooccur is determined as that pump set point speed where levels and phaselags between pump outlet and inlet pressures fall during diastole causeflow reversal.

In a particular preferred form the pump speed at which regurgitation iscalculated to occur is calculated according to:Nregurg=N(t) for Qdiastole=0L/min

BRIEF DESCRIPTION OF THE DRAWINGS

Embodiments of the present invention will now be described, withreference to the accompanying drawings, wherein.

FIG. 1 is a longitudinal cross-sectional view of a preferred embodimentof the invention;

FIG. 2 is a cross-sectional view taken generally along the line Z-Z ofFIG. 1;

FIG. 3A is a cross-sectional view of an impeller blade taken generallyalong the line A-A of FIG. 2;

FIG. 3B is an enlargement of the blade-pump housing interface portion ofFIG. 3A;

FIG. 3C is an alternative impeller blade shape;

FIGS. 4A, B, C illustrate various possible locations of magnet materialwithin a blade;

FIGS. 5A, B and C are left-hand end views of possible winding geometriestaken generally along the line S-S of FIG. 1;

FIG. 6 is a diagrammatic cross-sectional view of an alternativeembodiment of the invention as an axial pump;

FIG. 7 is an exploded, perspective view of a centrifugal pump assemblyaccording to a further embodiment of the invention;

FIG. 8 is a perspective view of the impeller of the assembly of FIG. 7;

FIG. 9 is a perspective, cut away view of the impeller of FIG. 8 withinthe pump assembly of FIG. 7;

FIG. 10 is a side section indicative view of the impeller of FIG. 8;

FIG. 11 is a detailed view in side section of blade portions of theimpeller of FIG. 10;

FIG. 12 is a block diagram of an electronic driver circuit for the pumpassembly of FIG. 7;

FIG. 13 is a graph of head versus flow for the pump assembly of FIG. 7;

FIG. 14 is a graph of pump efficiency versus flow for the pump assemblyof FIG. 7;

FIG. 15 is a graph of electrical power consumption versus flow for thepump assembly of FIG. 7;

FIG. 16 is a plan, section view of the pump assembly showing a volutearrangement according to a preferred embodiment;

FIG. 17 is a plan, section view of a pump assembly showing analternative volute arrangement;

FIG. 18 is a plan view of an impeller according to a further embodimentof the invention;

FIG. 19 is a plan view of an impeller according to a further embodimentof the invention;

FIG. 20 is a perspective view of an impeller according to a furtherembodiment of the invention;

FIG. 21 is a perspective view of an impeller according to yet a furtherembodiment of the invention;

FIG. 22 is a perspective, partially cut away view of an impelleraccording to yet a further embodiment of the invention;

FIG. 23 is a top, perspective view of the impeller of FIG. 22;

FIG. 24 is a perspective view of the impeller of FIG. 22 with its topshroud removed;

FIG. 25 illustrates an alternative embodiment wherein the deformedsurface is located on the pump housing; and

FIG. 26 illustrates a further embodiment wherein deformed surfaces arelocated both on the impeller and on the housing.

FIG. 27 illustrates diagrammatically the basis of operation of the“deformed surfaces” utilised for hydrodynamic suspension of embodimentsof the invention.

FIG. 28 is a block diagram of a non-contact estimation and controlsystem in accordance with a first embodiment of the invention applied toa blood pump;

FIG. 29 is a characteristic estimation curve utilised by the non-contactestimation and control system of FIG. 1;

FIG. 30 is a side, cut-away view of the pump of FIG. 1;

FIG. 31 is a plan, cut-away view of the coil and magnet system of thepump of FIG. 1;

FIG. 32 is a block diagram of an electronic driver circuit for the pumpassembly of FIG. 7;

FIG. 33 illustrates efficiency versus specific speed for a range of pumptypes, to be contrasted with the flat HQ curves of FIG. 13;

FIG. 34 provides a graphical comparison of HQ curves for pumpconstructions according to embodiments of the invention compared withtypical centrifugal pump HQ curves;

FIG. 35 illustrates a particular preferred form of impeller describedwith reference to example 2;

FIG. 36 illustrates an implanted rotary pump assembly and associatedcontrol system according to example 2;

FIG. 37 illustrates graphically a control strategy to avoid over pumpingfor the system of example 2;

FIG. 38 is a graphical illustration of application of the algorithms ofthe control system to estimate pressure head for example 2; and

FIG. 39 is a graphical illustration of application of the algorithms ofthe control system to provide a flow rate estimate for the system ofexample 2.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT

The pump assemblies according to various preferred embodiments to bedescribed below all have particular, although not exclusive, applicationfor implantation in a mammalian body so as to at least assist, if nottake over, the function of the mammalian heart. In practice this isperformed by placing the pump assembly entirely within the body of themammal and connecting the pump between the left ventricle and the aortaso as to assist left side heart function. It may also be connected tothe right ventricle and pulmonary artery to assist the right side of theheart.

In this instance the pump assembly includes an impeller which is fullysealed within the pump body and so does not require a shaft extendingthrough the pump body to support it. The impeller is suspended, in use,within the pump body by the operation of hydrodynamic forces imparted asa result of the interaction between the rotating impeller, the internalpump walls and the fluid which the impeller causes to be urged from aninlet of the pump assembly to an outlet thereof.

A preferred embodiment of the invention is the centrifugal pump 1, asdepicted in FIGS. 1 and 2, intended for implantation into a human, inwhich case the fluid referred to below is blood. The pump housing 2, canbe fabricated in two parts, a front part 3 in the form of a housing bodyand a back part 4 in the form of a housing cover, with a smooth jointherebetween, for example at 5 in FIG. 1. The pump 1 has an axial inlet6 and a tangential outlet 7. The rotating part 100 is of very simpleform, comprising only blades 8 and a blade support 9 to hold thoseblades fixed relative to each other. The blades may be curved asdepicted in FIG. 2, or straight, in which case they can be either radialor back-swept, i.e. at an angle to the radius. This rotating part 100will hereafter be called the impeller 100, but it also serves as abearing component and as the rotor of a motor configuration as to befurther described below whereby a torque is applied by electromagneticmeans to the impeller 100. Note that the impeller has no shaft and thatthe fluid enters the impeller from the region of its axis RR. Some ofthe fluid passes in front of the support 9 and some behind it, so thatthe pump 1 can be considered of two-sided open type, as compared toconventional open centrifugal pumps, which are only open on the frontside. Approximate dimensions found adequate for the pump 1 to perform asa ventricular assist device, when operating at speeds in the range 1,500rpm to 4,000 rpm, are outer blade diameter 40 mm, outer housing averagediameter 60 mm, and housing axial length 40 mm.

As the blades 8 move within the housing, some of the fluid passesthrough the gaps, much exaggerated in FIGS. 1 and 3, between the bladebearing faces 101 and the housing front face 10 and housing back face11. In all open centrifugal pumps, the gaps are made small because thisleakage flow lowers the pump hydrodynamic efficiency. In the pumpdisclosed in this embodiment, the gaps are made smaller than isconventional in order that the leakage flow can be utilised to create ahydrodynamic bearing. For the hydrodynamic forces to be sufficient, theblades may also be tapered as depicted in FIGS. 3A and 3B, so that thegap 104 is larger at the leading edge 102 of the blade 8 than at thetrailing edge 103 thereby providing one example of a wedge-shapedrestriction defined by at least one “deformed surface” as describedelsewhere in this specification and a corresponding opposing surface.The fluid 105 which passes through the gap thus experiences a wedgeshaped restriction which generates a thrust, as described in Reynolds'theory of lubrication (see, for example, “Modern Fluid Dynamics, Vol. 1Incompressible Flow”, by N. Curle and H. J. Davies, Van Nostrand, 1968).For blades considerably thinner than their length, the thrust isproportional to the square of the blade thickness at the bearing face,and thus in this embodiment thick blades are favoured, since if theproportion of the pump cavity filled by blades is constant, then the netthrust force will be inversely proportional to the number of blades.However, the blade bearing faces can be made to extend as tails fromthin blades as depicted in FIG. 3C in order to increase the blade facearea adjacent the walls.

In one particular form, the tails join adjacent blades so as to form acomplete shroud with wedges or tapers incorporated therein. An exampleof a shroud design as well as other variations on the blade structurewill be described later in this specification.

For manufacturing simplicity, the housing front face 10 can be madeconical, with an angle of around 450 so that it provides both axial andradial hydrodynamic forces. Other angles are suitable that achieve thefunctional requirements of this pump including the requirements for bothaxial and radial hydrodynamic forces.

Other curved surfaces are possible provided both axial and radialhydrodynamic forces can be produced as a result of rotation of theblades relative to the housing surfaces.

In one form the housing back face 11 may include a roughly conicalextension 12 pointing into the pump cavity 106, to eliminate or minimisethe effect of the flow stagnation point on the axis of the back housing.

Alternatively extension 12 can resemble an impeller eye to make the flowmixed.

In an alternative form the extension 12 can be omitted for ease ofmanufacture.

In this preferred embodiment, for manufacturing simplicity and foruniformity in the flow axial direction RR, the housing back face 11 ismade flat over the bearing surfaces, i.e. under the blade bearing faces.With this the case, a slacker tolerance on the alignment between theaxes of the front part 3 and back part 4 of the housing 2 ispermissible. An alternative is to make the back face 11 conical at thebearing surfaces, with taper in the opposite direction to the front face10, so that the hydrodynamic forces from the back face will also haveradial components. Tighter tolerance on the axes alignment would then berequired, and some of the flow would have to undergo a reversal in itsaxial direction.

There are many profiles of bearing surface which will generate thewedge-shaped restriction. In the preferred embodiment the amount ofmaterial removed simply varies linearly or approximately linearly acrossthe blade between the body and trailing edges. Alternative taper shapescan include a radiused leading edge or a step in the blade bearing face,though the corner in that step may represent a stagnation line posing athrombosis risk.

For a given minimum gap, at the trailing blade edge, the hydrodynamicforce is maximal if the gap at the leading edge of the blade end face isapproximately double that at the trailing edge of the blade end face.Thus the taper, which equals the blade face leading edge gap minus thetrailing edge gap, should be chosen to match a nominal minimum gap, oncethe impeller has shifted towards that edge. Dimensions which have beenfound to give adequate thrust forces are a taper of around 0.05 mm for anominal minimum gap of around 0.05 mm, and an average circumferentialblade bearing face thickness of around 60 mm for 4 blades. For the frontface, the taper is measured within the plane perpendicular to the axis.The axial length of the housing between the front and back faces at anyposition should then be made about 0.2 mm greater than the axial lengthof the blade, when it is coaxial with the housing, so that the minimumgaps are both about 0.1 mm axially when the impeller 100 is centrallypositioned within the housing 2. Then, for example, if the impellershifts axially by 0.05 mm, the minimum gaps will be 0.05 mm at one faceand 0.15 mm at the other face. The thrust increases with decreasing gapand would be much larger from the 0.05 mm gap than from the 0.15 mm gap,about 14 times larger for the above dimensions. Thus there is a netrestoring force away from the smaller gap.

Similarly, for radial shifts of the impeller the radial component of thethrust from the smaller gap on the conical housing front face wouldoffer the required restoring radial force. The axial component of thatforce and its torque on the impeller would have to be balanced by anaxial force and torque from the housing back face, and so the impellerwill also have to shift axially and tilt its axis to be no longerparallel with the housing axis. Thus as the person moves and the pump isaccelerated by external forces, the impeller will continually shift itsposition and alignment, varying the gaps in such a way that the totalforce and torque on the impeller 100 match that demanded by inertia. Thegaps are so small, however, that the variation in hydrodynamicefficiency will be small, and the pumping action of the blades will beapproximately the same as when the impeller is centrally located.

While smaller gaps imply greater hydrodynamic efficiency and greaterbearing thrust forces, smaller gaps also demand tighter manufacturingtolerances, increase frictional drag on the impeller, and impose greatershear stress an the fluid. Taking these points in turn, for the above0.05 mm tapers and gaps, tolerances of around 0.005 mm are needed, whichimposes some cost penalty but is achievable. A tighter tolerance isdifficult, especially if the housing is made of a plastic, given thechanges in dimension caused by temperature and possible absorption offluid by plastic materials which may be in contact with the blood suchas Acrylic of polyurethane. The frictional drag for the above gapsproduces much smaller torque than the typical motor torque. Finally, toestimate the shear stress, consider a rotation speed of 3,000 rpm and atypical radius of 15 mm, at which the blade speed is 4.7 ms⁻¹ and theaverage velocity shear for an average gap of 0.075 mm is 6.2×10⁴ s⁻¹.For blood of dynamic viscosity 3.5×10⁻³ kgm-¹s-¹, the average shearstress would be 220 Nm⁻². Other prototype centrifugal blood pumps withclosed blades have found that slightly larger gaps, e.g. 0.15 mm, areacceptable for haemolysis. A major advantage of the open blades of thepresent invention is that a fluid element that does pass through a bladebearing face gap will have very short residence time in that gap, around2×10⁻³ s, and the fluid element will most likely be swept though thepump without passing another blade bearing face.

With particular reference to FIGS. 3A and 3B typical working clearancesand working movement for the impeller 8 with respect to the upper andlower housing surfaces 10, 11 is of the order of 100 microns clearanceat the top and at the bottom. In use gravitational and other forces willbias the impeller 8 closer to one or other of the housing wallsresulting, typically in a clearance at one interface of the order of 50microns and a corresponding larger clearance at the other interface ofthe order of 150 microns. In use, likely maximum practical clearanceswill range from 300 microns down to 1 micron.

Typical restoring forces for a 25 gram rotor mass spinning at 2200 rpmare 1.96 Newtons at a 20 micron clearance extending to 0.1 Newtons at an80 micron clearance.

To minimise the net force required of the hydrodynamic bearings, the netaxial and radial hydrodynamic forces on the impeller from the bulk fluidflow should be minimised, where “bulk” here means other than from thebearing thrust surfaces.

The radial force on the impeller depends critically on the shape of theoutput flow collector or volute 13. The shape should be designed tominimise the radial impeller force over the desired range of pumpspeeds, without excessively lowering the pump efficiency. The optimalshape will have a roughly helical perimeter between the “cutwater” andoutlet. The radial force can also be reduced by the introduction of aninternal division in the volute 13 to create a second output flowcollector passage, with tongue approximately diametrically opposite tothe tongue of the first passage.

An indicative plan view of impeller 100 relative to housing 2 is shownin FIG. 2 having a concentric volute 13.

FIG. 17 illustrates the alternative volute arrangement comprising asplit volute created by volute barrier 107 which causes volute 108 in afirst hemisphere of the housing 2 to split into first half volute 109and second half volute 110 over the second hemisphere. The hemispheresare defined respectively on each side of a diameter of the housing 2which passes through or near exit point 111 of outlet 7.

In alternative forms concentric volutes can be utilised, particularlywhere specific speed is relatively low.

In a further particular form a vaneless diffuser may also reduce theradial force.

In regard to the bulk hydrodynamic axial force, if the bladecross-section is made uniform in the axial direction along therotational axis, apart from the conical front edge surface, then thepressure acting on the blade surface (excluding the bearing surfaces)will have no axial component. This also simplifies the blademanufacture. The blade support 9 should then be shaped to minimise axialthrust on the impeller and minimise disturbance to the flow over therange of speeds, while maintaining sufficient strength to preventrelative blade movement. The key design parameter affecting the axialforce is the angle of the support. The support is drawn in FIG. 1 ashaving the same internal diameter as the blades, which may aidmanufacture. However, the support could be made with larger or smallerinternal diameter to the blades. There may be advantage in using anon-axisymmetric support, e.g. with larger radius on the trailingsurface of a blade than the radius at the leading surface of the nextblade. If the blades are made with non-uniform cross-section to increasehydrodynamic efficiency, then any bulk hydrodynamic axial force on themcan be balanced by shaping the support to produce an opposite bulkhydrodynamic axial force on it.

Alternatively, by careful manufacture of taper axial thrust can beengineered.

Careful design of the entire pump, employing computational fluiddynamics, is necessary to determine the optimal shapes of the blades 8,the volute 13, the support and the housing 2, in order to maximisehydrodynamic efficiency while keeping the bulk fluid hydrodynamicforces, shear and residence times low. All edges and the joins betweenthe blades and the support should be smoothed.

The means of providing the driving torque on the impeller 100 of thepreferred embodiment of the invention is to encapsulate permanentmagnets 14 in the blades 8 of the impeller 100 and to drive them with arotating magnetic field pattern from oscillating currents in windings 15and 16, fixed relative to the housing 2. Magnets of high remanence suchas sintered rare-earth magnets should be used to maximise motorefficiency. The magnets can be aligned axially but greater motorefficiency is achieved by tilting the magnetisation direction to anangle of around 15° to 30° outwards from the inlet axis, with 22.5° tiltsuitable for a body of conical angle 45°. The magnetisation directionmust alternate in polarity for adjacent blades. Thus there must be aneven number of blades. Since low blade number is preferred for thebearing force, and since two blades would not have sufficient bearingstiffness to rotation about an axis through the blades and perpendicularto the pump housing (unless the blades are very curved), four blades arerecommended. A higher number of blades, for example 6 or 8 will alsowork.

Some possible options for locating the magnets 14 within the blades 8are shown in FIG. 4. The most preferred which is depicted in FIG. 4A, isfor the blade to be made of magnet material apart from a biocompatibleshell or coating to prevent fluid corroding the magnets and to preventmagnet material (which may be toxic) entering the blood stream. Thecoating should also be sufficiently durable especially at blade cornersto withstand rubbing during start-up or during inadvertent bearing touchdown.

In one particular form the inside walls of the pump housing 2 are alsocoated with a biologically compatible and wear resistant material suchas titanium nitride so that wear on both of the touching surfaces isminimised.

An acceptable coating thickness is approximately 1 micron.

In one form the magnet material can be potted in titanium or a polymerichousing which is then, in turn, coated with a biologically compatibleand tough material such as titanium nitride.

In an alternative form a suitable impeller manufacturing method is todie-press the entire impeller, blades and support, as a single axiallyaligned magnet. The die-pressing is much simplified if near axiallyuniform blades are used (blades with an overhang such as in FIG. 3C areprecluded). During pressing, the crushed rare-earth particles must bealigned in an axial magnetic field. This method of die-pressing withparallel alignment direction is cheaper for rare-earth magnets, althoughit produces slightly lower remanence magnets. The tolerance indie-pressing is poor, and grinding of the tapered blade surfaces isrequired. Then the magnet impeller can be coated, for example byphysical vapour deposition, of titanium nitride for example, or bychemical vapour deposition, of a teflon coating.

Finally, to create the alternating blade polarity the impeller may beplaced in a special pulse magnetisation fixture, with an individual coilsurrounding each blade. The support of a die-pressed magnet impelleracquires some magnetisation near the blades, with negligible influence.

Alternative magnet locations are sketched in FIG. 4B and FIG. 4C inwhich quadrilateral or circular cross-section magnets 14 are insertedinto the blades. Sealing and smoothing of the blade bearing surfacesover the insertion holes is then required to reinstate the taper.

All edges in the pump should be radiused and surfaces smoothed to avoidpossible damage to formed elements of the blood.

The windings 15 and 16 of the preferred embodiment are slotless orair-gap windings with the same pole number as the impeller, namely fourpoles in the preferred embodiment. A ferromagnetic iron yoke 17 ofconical form for the front winding and an iron ferromagnetic yoke 18 ofannular form for the back winding may be placed on the outside of thewindings to increase the magnetic flux densities and hence increasemotor efficiency. The winding thicknesses should be designed for maximummotor efficiency, with the sum of their axial thicknesses somewhat lessthan but comparable to the magnet axial length. The yokes can be made ofsolid ferromagnetic material such as iron. To reduce “iron” losses, theyokes 17 can be laminated, for example in layers or by helically windingthin strip, or can be made of iron/powder epoxy composite. The yokesshould be positioned such that there is zero net axial magnetic force onthe impeller when it is positioned centrally in the housing. Themagnetic force is unstable and increases linearly with axialdisplacement of the impeller away from the central position, with thegradient being called the negative stiffness of the magnetic force. Thisunstable magnetic force must be countered by the hydrodynamic bearings,and so the stiffness should be made as small as possible. Choosing theyoke thickness such that the flux density is at the saturation levelreduces the stiffness and gives minimum mass. An alternative can be tohave no iron yokes, completely eliminating the unstable axial magneticforce, but the efficiency of such designs may be lower and the magneticflux density in the immediate vicinity of the pump may violate safetystandards and produce some tissue heating. In any case, the stiffness isacceptably small for slotless windings with the yokes present. Anotheralternative would be to insert the windings in slots in laminated ironstators which would increase motor efficiency and enable use of lessmagnet material and potentially lighter impeller blades. However, theunstable magnetic forces would be significant for such slotted motors.Also, the necessity for fat blades to generate the required bearingforces in this embodiment allows room for large magnets, and so slotlesswindings are chosen in the preferred embodiment.

Instead of determining the yoke positions so that the impeller has zeromagnetic axial force in the central position, it may be possible toprovide a bias axial magnetic force on the impeller, which cancounteract other forces such as any average bulk hydrodynamic axialforce. In particular, by ensuring a net axial force into the conicalbody, the thrust bearings on the cover surface can be made superfluous.However, such a bias would demand greater average thrust forces, smallergaps and increased blood damage, and so the recommended goal is to zeroboth the magnetic and bulk hydrodynamic axial forces on the impellerwhen centrally positioned.

The overall design requirement for exclusive hydrodynamic suspensionrequires control of the external force balance to make the relativemagnitude of hydrodynamic thrust sufficient to overcome the externalforces. Typical external forces include gravitational forces and netmagnetic forces arising as a result of the motor drive.

There are many options for the winding topology and number of phases.FIG. 5A depicts the preferred topology for the body winding 15, viewedfrom the inlet axis.

The cover winding 16 looks similar but the coils need not avoid theinlet tube and so they appear more triangular in shape. The body windinghas a more complex three dimensional shape with bends at the ends of thebody support section. Each winding consists of three coils. Each coil ismade from a number of turns of an insulated conductor such as copperwith the number of turns chosen to suit the desired voltage. The coilside mid-lines span an angle of about 50°-100° at the axis when thecoils are in position. The coils for body and cover are aligned axiallyand the axially adjacent coils are connected in either parallel orseries connection to form one phase of the three phase winding. Parallelconnection offers one means of redundancy in that if one coil fails, thephase can still carry current through the other coil. In parallelconnection each of the coil and body winding has a neutral pointconnection as depicted in FIG. 5A, whereas in series connection, onlyone of the windings has a neutral point.

An alternative three phase winding topology, depicted in FIG. 5S, usesfour coils per phase for each of the body and cover windings, with eachcoil wrapping around the yoke, a topology called a “Gramm ring” winding.

Yet another three phase winding topology, depicted in FIG. 5C, uses twocoils per phase for each of the body and cover windings, and connectsthe coil sides by azimuthal end-windings as is standard motor windingpractice. The coils are shown tilted to approximately follow the bladecurvature, which can increase motor efficiency, especially for the phaseenergising strategy to be described below in which only one phase isenergised at a time. The winding construction can be simplified bylaying the coils around pins protruding from a temporary former, thepins shown as dots in 2 rings of 6 pins each in FIG. 5C. The coils arelabeled alphabetically in the order in which they would be layed, coilsa and d for phase A, b and e for phase B, and c and f for phase C.Instead of or as well as pins, the coil locations could be defined bythin fins, running between the pins in FIG. 5C, along the boundarybetween the coils. The coil connections depicted in FIG. 5C are thoseappropriate for the winding nearest the motor terminals for the case ofseries connection, with the optional lead from the neutral point on theother winding included.

The winding topologies depicted in FIGS. 5B and C allow the possibilityof higher motor efficiency but only if significantly higher coil mass isallowed, and since option FIG. 5A is more compact and simpler tomanufacture, it is the preferred option. Material ribs between the coilsof option FIG. 5A can be used to stiffen the housing.

Multi-stranded flexible conductors within a suitable biocompatible cablecan be used to connect the motor windings to a motor controller. Theenergisation of the three phases can be performed by a standardsensorless controller, in which two out of six semiconducting switchesin a three phase bridge are turned on at any one time. Alternatively,because of the relatively small fraction of the impeller cross-sectionoccupied by magnets, it may be slightly more efficient to only activateone of the three phases at a time, and to return the current by aconductor from the neutral point in the motor. Careful attention must bepaid to ensure that the integrity of all conductors and connections isfailsafe.

In one embodiment, the two housing components 3 and 4 are made byinjection moulding from non-electrically conducting plastic materialssuch as Lexan polycarbonate plastic. Alternatively the housingcomponents can be made from ceramics. The windings and yokes are ideallyencapsulated within the housing during fabrication moulding. In thisway, the separation between the winding and the magnets is minimised,increasing the motor efficiency, and the housing is thick, increasingits mechanical stiffness. Alternatively, the windings can be positionedoutside the housing, of thickness at least around 2 mm for sufficientstiffness.

If the housing material plastic is hygroscopic or if the windings areoutside the housing, it may be necessary to first enclose the windingsand yoke in a very thin impermeable shell. Ideally the shell should benon-conducting (such as ceramic or plastic). Titanium of around 0.1 mmto 0.2 mm thickness gives sufficiently low eddy losses. Encapsulationwithin such a shell is needed to prevent winding movement.

Alternatively, and in a particularly preferred embodiment the housingcomponents 3 and 4 may be made from a biocompatible metallic material oflow electrical conductivity, such as Ti-6A1-4V. To minimise the eddycurrent loss, the material must be as thin as possible, e.g. 0.1 mm to0.5 mm, wherever the material experiences high alternating magnetic fluxdensities, such as between the coils and the housing inner surfaces 10and 11.

The combining of the motor and bearing components into the impeller inthe preferred embodiment provides several key advantages. The rotorconsequently has very simple form, with the only cost of the bearingbeing tight manufacturing tolerances. The rotor mass is very low,minimising the bearing force needed to overcome weight. Also, with thebearings and the motor in the same region of the rotor, the bearingsforces are smaller than if they had to provide a torque to supportmagnets at an extremity of the rotor.

A disadvantage of the combination of functions in the impeller is thatits design is a coupled problem. The optimisation should ideally linkthe fluid dynamics, magnetics and bearing thrust calculations. Inreality, the blade thickness can be first roughly sized to give adequatemotor efficiency and sufficient bearing forces with a safety margin.Fortuitously, both requirements are met for four blades of approximateaverage circumferential thickness 6 mm or more. The housing, blade, andsupport shapes can then be designed using computational fluid dynamics,maintaining the above minimum average blade thickness. Finally the motorstator, i.e. winding and yoke, can be optimised for maximum motorefficiency.

FIG. 6 depicts an alternative embodiment of the invention as an axialpump. The pump housing is made of two parts, a front part 19 and a backpart 20, joined for example at 21. The pump has an axial inlet 22 andaxial outlet 23. The impeller comprises only blades 24 mounted on asupport cylinder 25 of reducing radius at each end. An important featureof this embodiment is that the blade bearing surfaces are tapered togenerate hydrodynamic thrust forces which suspend the impeller. Theseforces could be used for radial suspension alone from the straightsection 26 of the housing, with some alternative means used for axialsuspension, such as stable axial magnetic forces or a conventionaltapered-land type hydrodynamic thrust bearing. FIG. 6 proposes a designwhich uses the tapered blade bearing surfaces to also provide an axialhydrodynamic bearing. The housing is made with a reducing radius at itsends to form a front face 27 and a back face from which the axialthrusts can suspend the motor axially. Magnets are embedded in theblades with blades having alternating polarity and four blades beingrecommended. Iron in the outer radius of the support cylinder 25 can beused to increase the magnet flux density. Alternatively, the magnetscould be housed in the support cylinder and iron could be used in theblades. A slotless helical winding 29 is recommended, with outwardbending end-windings 30 at one end to enable insertion of the impellerand inward bending windings 31 at the other end to enable insertion ofthe winding into a cylindrical magnetic yoke 32. The winding can beencapsulated in the back housing part 20.

Third Embodiment

With reference to FIGS. 7 to 15 inclusive there is shown a furtherpreferred embodiment of the pump assembly 200.

With particular reference initially to FIG. 7 the pump assembly 200comprises a housing body 201 adapted for bolted connection to a housingcover 202 and so as to define a centrifugal pump cavity 203 therewithin.

The cavity 203 houses an impeller 204 adapted to receive magnets 205within cavities 206 defined within blades 207. As for the firstembodiment the blades 207 are supported from a support 208.

Exterior to the cavity 203 but forming part of the pump assembly 200there is located a body winding 209 symmetrically mounted around inlet210 and housed between the housing body 201 and a body yoke 211.

Also forming part of the pump assembly 200 and also mounted external topump cavity 203 is cover winding 212 located within winding cavity 213which, in turn, is located within housing cover 202 and closed by coveryoke 214.

The windings 212 and 209 are supplied from the electronic controller ofFIG. 12 as for the first embodiment the windings are arranged to receivea three phase electrical supply and so as to set up a rotating magneticfield within cavity 203 which exerts a torque on magnets 205 within theimpeller 204 so as to urge the impeller 204 to rotate substantiallyabout central axis TT of cavity 203 and in line with the longitudinalaxis of inlet 210. The impeller 204 is caused to rotate so as to urgefluid (in this case blood) around volute 215 and through outlet 216.

The assembly is bolted together in the manner indicated by screws 217.The yokes 211, 214 are held in place by fasteners 218. Alternatively,press fitting is possible provided sufficient integrity of seal can bemaintained.

In a particularly preferred form the components are welded together.

FIG. 8 shows the impeller 204 of this embodiment and clearly shows thesupport 208 from which the blades 207 extend. The axial cavity 219 whichis arranged, in use, to be aligned with the longitudinal axis of inlet210 and through which blood is received for urging by blades 207 isclearly visible.

The cutaway view of FIG. 9 shows the axial cavity 219 and also themagnet cavities 206 located within each blade 207. The support structure220 extending from housing cover 202 aligned with the axis of inlet 210and axial cavity 219 of impeller 204 is also shown.

FIG. 10 is a side section, indicative view of the impeller 204 definingthe orientations of central axis FF, top taper face DD and bottom taperface BB, which tapers are illustrated in FIG. 11 in side section view.

FIG. 11A is a section of a blade 207 of impeller 204 taken through planeDD as defined in FIG. 10 and shows the top edge surface 221 to beprofiled from a leading edge 223 to a trailing edge 224 as follows:central portion 227 comprises an ellipse with centre on the dashedmidline having a semi-major axis of radius 113 mm and a semi-minor axisof radius 80 mm and then followed by leading conical surface 225 andtrailing conical surface 226 on either side thereof as illustrated inFIG. 11A. The leading surface 225 has radius 0.05 mm less than thetrailing surface 226. This prescription is for a taper which can beachieved by a grinding wheel, but many alternative prescriptions couldbe devised to give a taper of similar utility.

The leading edge 223 is radiused as illustrated.

FIG. 11B illustrates in cross-section the bottom edge face 222 of blade207 cut along plane BB of FIG. 10.

The bottom face includes cap 228 utilised for sealing magnet 205 withincavity 206.

In this instance substantially the entire face comprises a straighttaper with a radius of 0.05 mm at leading edge 229 and a radius of 0.25mm at trailing edge 230.

The blade 207 is 6.0 mm in width excluding the radii at either end.

FIG. 12 comprises a block diagram of the electrical controller suitablefor driving the pump assembly 200 and comprises a three phasecommutation controller 232 adapted to drive the windings 209, 212 of thepump assembly. The commutation controller 232 determines relative phaseand frequency values for driving the windings with reference to setpoint speed input 233 derived from physiological controller 234 which,in turn, receives control inputs 235 comprising motor current input andmotor speed (derived from the commutation controller 232). Whilst notpreferred, patient blood flow 236, and venous oxygen saturation 237 canbe input as well. The pump blood flow can be approximately inferred fromthe motor speed and current via curve-fitted formulae.

FIG. 13 is a graph of pressure against flow for the pump assembly 200where the fluid pumped is 18% glycerol for impeller rotation velocityover the range 1500 RPM to 2500 RPM. The 18% glycerol liquid is believedto be a good analogue for blood under certain circumstances, for examplein the housing gap.

FIG. 14 graphs pump efficiency against flow for the same fluid over thesame speed ranges as for FIG. 13.

FIG. 15 is a graph of electrical power consumption against flow for thesame fluid over the same speed ranges as for FIG. 13.

The common theme running through the first, second and third embodimentsdescribed thus far is the inclusion in the impeller of a taper or otherdeformed surface which, in use, moves relative to the adjacent housingwall thereby to cause a restriction with respect to the line of movementof the taper or deformity thereby to generate thrust upon the impellerwhich includes a component substantially normal to the line of movementof the surface and also normal to the adjacent internal pump wall withrespect to which the restriction is defined for fluid locatedtherebetween.

In order to provide both radial and axial direction control at least oneset of surfaces must be angled with respect to the longitudinal axis ofthe impeller (preferably at approximately 45° thereto) thereby togenerate or resolve opposed radial forces and an axial force which canbe balanced by a corresponding axial force generated by at least oneother tapered or deformed surface located elsewhere on the impeller.

In the forms thus far described top surfaces of the blades 8, 207 areangled at approximately 45° with respect to the longitudinal axis of theimpeller 100, 204 and arranged for rotation with respect to the internalwalls of a similarly angled conical pump housing. The top surfaces ofthe blades are deformed so as to create the necessary restriction in thegap between the top surfaces of the 7 blades and the internal walls ofthe conical pump housing thereby to generate a thrust which can beresolved to both radial and axial components.

In the examples thus far the bottom faces of the blades 8, 207 comprisesurfaces substantially lying in a plane at right angles to the axis ofrotation of the impeller and, with their deformities define a gap withrespect to a lower inside face of the pump housing against which asubstantially only axial thrust is generated.

Other arrangements are possible which will also, relying on theseprinciples, provide the necessary balanced radial and axial forces. Sucharrangements can include a double support arrangement where the conicaltop surface of the blades is mirrored in a corresponding bottom conicalsurface. The only concern with this arrangement is the increased depthof pump which can be a problem for in vivo applications where sizeminimisation is an important criteria.

Fourth Embodiment

With reference to FIG. 18 a further embodiment of the invention isillustrated comprising a plan view of the impeller 300 forming part of a“channel” pump. In this embodiment the blades 301 have been widenedrelative to the blades 207 of the third embodiment to the point wherethey are almost sector-shaped and the flow gaps between adjacent blades301, as a result, take the form of a channel 302, all in communicationwith axial cavity 303.

A further modification of this arrangement is illustrated in FIG. 19wherein impeller 304 includes secter-shaped blades 305 having curvedleading and trailing −38 portions 306, 307 respectively thereby definingchannels 308 having fluted exit portions 309.

As with the first and second embodiments the radial and axialhydrodynamic forces are generated by appropriate profiling of the topand bottom faces of the blades 301, 305 (not shown in FIGS. 18 and 19).

FIG. 20 illustrates a perspective view of an impeller 304 which followsthe theme of the impeller arrangement of FIGS. 18 and 19 in perspectiveview and where like parts are numbered as for FIG. 19. In this case thefour blades 305 are joined at mid-portions thereof by a blade support inthe form of a conical rim 350 and have face portions which are shaped soas to have an increased curvature on the pressure face 351 thereofcompared with the suction face 352.

Fifth Embodiment

A fifth embodiment of a pump assembly according to the inventioncomprises an impeller 410 as illustrated in FIG. 21 where, conceptually,the upper and lower surfaces of the blades of previous embodiments areinterconnected by a top shroud 411 and a bottom shroud 412. In thisembodiment the blades 413 can be reduced to a very small width as thehydrodynamic behaviour imparted by their surfaces in previousembodiments is now given effect by the profiling of the shrouds 411, 412each of which, in this instance, comprise a series of smoothed wedges414 with the leading edge of one wedge directly interconnected to thetrailing edge of the preceding wedge.

As for previous embodiments the top shroud 411 is of overall conicalshape thereby to impart both radial and axial thrust forces whilst thebottom shroud 412 is substantially planar thereby to impartsubstantially only axial thrust forces.

It is to be understood that, whilst the example of FIG. 21 shows thesurfaces of the shroud 411 angled at approximately 450 to the vertical,other inclinations are possible extending to an inclination of 00 to thevertical which is to say the impeller 410 can take the form of acylinder with surface rippling or other deformations which impart thenecessary hydrodynamic lift, in use.

With reference to FIGS. 22 to 24 a specific example of the conceptembodied in FIG. 21 is illustrated and wherein like components arenumbered as for FIG. 21.

It will be observed that, with reference to FIG. 24, the blades 413 arethin compared to previous embodiments and, in this instance, are arcuatechannels 416 therebetween which allow fluid communication from a centrevolume 417 to the periphery 418 of the impeller 410.

In this arrangement it will be noted that the wedges 414 are separatedone from the other on each shroud by channels 419. The channels extendradially down the shroud from the centre volume 417 to the periphery418.

In such designs with thin blades, the magnets required for the drivingtorque can be contained within the top or bottom shroud-or both, alongwith the optional soft magnetic yokes to increase motor efficiency.

A variation of this embodiment is to have the wedge profiling cut intothe inner surfaces of the housing and have smooth shroud surfaces.

Sixth Embodiment

In contrast to the embodiments illustrated with respect to FIGS. 3A, 3Band 3C an arrangement is shown in FIG. 25 wherein the “deformed surface”comprises a stepped formation 510 forming part of an inner wall of thepump housing (not shown). In this instance the rotor including blade 511includes a flat working surface 512 (and not having a deformed surfacetherein) which is adapted for relative movement in the direction of thearrow shown with respect to the stepped formation 510 thereby togenerate hydrodynamic thrust therebetween.

Seventh Embodiment

With reference to FIG. 26 there is shown an arrangement having facingdeformed surfaces. The rotor blade 610 includes a deformed surface 612at a working face thereof. In this instance the deformation comprisescurved edge 613. Relative movement of the rotor blade 610 in thedirection of the arrow with respect to deformed facing surface 611forming part of the pump housing (not shown) causes relativehydrodynamic thrust therebetween.

The foregoing describes principles and examples of the presentinvention, and modifications, obvious to those skilled in the art, canbe made thereto without departing from the scope and spirit of theinvention.

Principles of Operation

With particular reference to FIG. 27 this specification describes thesuspension of an impeller 600 within a pump housing 601 by the use ofhydrodynamic forces. In this specification the suspension of theimpeller 600 is performed dominantly which is to say exclusively byhydrodynamic forces.

The hydrodynamic forces are forces which are created by relativemovement between two surfaces which have a fluid in the gap between thetwo surfaces. In the case of the use of the pump assembly 602 as arotary blood pump the fluid is blood.

The hydrodynamic forces can arise during relative movement between twosurfaces even where those surfaces are substantially entirely parallelto each other or non-deformed. However, in this specification,hydrodynamic forces are caused to arise during relative movement betweentwo surfaces where at least one of the surfaces includes a “deformedsurface”.

In this specification “deformed surface” means a surface which includesan irregularity relative to a surface which it faces such that, when thesurface moves in a predetermined direction relative to the surface whichit faces the fluid located in the gap therebetween experiences a changein relative distance between the surfaces along the line of movementthereby to cause a hydrodynamic force to arise therebetween in the formof a thrust force including at least a component substantially normal tothe plane of the gap defined at any given point between the facingsurfaces.

In the example of FIG. 27 there is a first deformed surface 603 formingat least part of a first face 604 of impeller 600 and a second deformedsurface 605 on a second face 606 of the impeller 600.

The inset of FIG. 27 illustrates conceptually how the first deformedsurface 603 may form only part of the first face 604.

The first deformed surface 603 faces first inner surface 607 of the pumphousing 601 whilst second deformed surface 605 faces second innersurface 608 of the pump housing 601.

In use first gap 609 defined between first deformed surface 603 andfirst inner surface 607 has a fluid comprising blood located thereinwhilst second gap 610 defined between second deformed surface 605 andsecond inner surface 608 also has a fluid comprising blood locatedtherein.

In use impeller 600 is caused to rotate about impeller axis 611 suchthat relative movement across first gap 609 between first deformedsurface 603 and first inner face 607 occurs and also relative movementacross second gap 610 between second deformed surface 605 and secondinner surface 608 occurs. The orientation of the deformities of firstdeformed surface 603 and second deformed surface 605 relative to theline of movement of the deformed surfaces 603, 605 relative to the innersurfaces 607, 608 is such that the fluid in the gaps 609, 610experiences a change in height of the gap 609, 610 as a function of timeand with the rate of change dependant on the shape of the deformities ofthe deformed surfaces and also the rate of rotation of the impeller 600relative to the housing 601. That is, at any given point on either innersurface 607 or 608, the height of the gap between the inner surface 607or 608 and corresponding deformed surface 603 or 605 will vary with timedue to passage of the deformed surface 603 or 605 over the innersurface.

Hydrodynamic forces in the form of thrust forces normal to the line ofrelative movement of the respective deformed surfaces 603, 605 relativeto the inner surfaces 607, 608 thus arise.

With this configuration it will be noted that the first gap 609 liessubstantially in a single plane whilst the second gap 610 is in the formof a support and angled at an acute angle relative to the plane of thefirst gap 609.

Accordingly, the thrust forces which can be enlisted to first gap 609and second gap 610 are substantially normal to and distributed acrossboth the predominantly flat plane of first deformed surface 603 andnormal to the substantially conical surface of second deformed surface605 thereby permitting restoring forces to be applied between theimpeller 600 and the pump housing 601 thereby to resist forces whichseek to translate the impeller 600 in space relative to the pump housing601 and also to rotate the impeller 600 about any axis (other than aboutthe impeller axis 611) relative to the pump housing 601. Thisarrangement substantially resists five degrees of freedom of movement ofimpeller 600 with respect to the housing 601 and does so predominantlywithout any external intervention to control the position of theimpeller with respect to the housing given that disturbing forces fromother sources, most notably magnetic forces on the impeller due to itsuse as rotor of the motor are net zero when the impeller occupies asuitable equilibrium position. The balance of all forces on the rotoreffected by manipulation of magnetic and other external sources may beadjusted such that the rotor is predominantly hydrodynamically born.

It will be observed that these forces increase as the gaps 609, 610narrow relative to a defined operating position and decrease as the gaps609, 610 increase relative to a defined operating gap. Because of theopposed orientation of first deformed surface 603 relative to seconddeformed surface 605 it is possible to design for an equilibriumposition of the impeller 600 within the pump housing 601 at a definedequilibrium gap distance for gaps 609, 610 at a specified rotorrotational speed about axis 611 and rotor mass leading to a closeapproximation to an unconditionally stable environment for the impeller600 within the pump housing 601 against a range of disturbing forces.

In this state the impeller 600 is effectively suspended exclusively byhydrodynamic thrust faces.

Characteristics and advantages which flow from the arrangement describedabove and with reference to the embodiments includes.

-   -   1. Low running speed, hence low haemolysis and controlled fluid        dynamics (especially shear stress) in the gap between the casing        and impeller. This in turn can lead to the selection of radial        off-flow and minimal incidence at on-flow to the rotor;    -   2. Radial or near-radial off-flow from the impeller can be        chosen in order to yield a “flat” pump characteristic (HQ)        curve.        Control System—Detailed Description

Embodiments of the present invention relate to a non-contact estimationand control system usable, although not exclusively, with blood PUMPsystems of the type illustrated in FIG. 28.

In this instance the estimation and control system 10 operates on andreceives sensor feedback from pump assembly adapted for implantation inhuman body 12 and arranged to operate in parallel across at least a partof heart 13 so as to at least assist if not fully take over the pumpingfunction of heart 13.

The pump assembly 11 includes an impeller 14 having vanes 15 which, whenurged to rotate by a magnetic field generated in one or more of coils16, 17 generates a pressure head H across the pump assembly 11 andcauses a flow of blood Q therethrough. In this instance the impeller 14is both a radial pump impeller and a rotor of motor 18 by virtue of theinclusion of magnets (not shown) within at least part of the impeller14.

Monitoring means 19 is adapted to sense electric current appearing inone or more of coils 16, 17 via sensing line 35 which, in conjunctionwith monitoring of voltage derived from commutation controller 32 (whichinjects current into one or more of the same coils 16, 17) permits themonitoring means 19 to derive power input (P_(in)) consumed by motor 18and actual rate of rotation of the motor/impeller 14 (n_(a)).

By means of equation 1.1 (in FIG. 28) it is thereby possible formonitoring means 19 to calculate an estimation of flow Q (and/or head H)for input into microprocessor 20. Microprocessor 20 accepts theseestimates and, together with other desired set points and predeterminedvalues calculates a desired set motor speed n_(set) which commutationcontroller 32 accepts via line 33. The commutation controller 32 theninjects current into one or more of coils 16, 17 in order to causeimpeller 14 to rotate at that set (desired) speed.

FIG. 20 illustrates the impeller 14 utilised in example 1 (to follow) ingreater detail.

FIG. 3 illustrates the head versus flow characteristic achievable withthe impeller of FIG. 20 for a number of different motor powers (Pin).FIG. 29 illustrates the characteristic curve used by the monitoringmeans 19 for example 1 (to follow) in accordance with the equation 1.1.

EXAMPLE 1

Flow rate and pressure difference (or head) are key variables needed inthe control of implantable rotary blood pumps. However, use of invasiveflow and/or pressure probes can decrease reliability and increase systempower consumption and expense. For given fluid viscosity, the flow stateis determined by any two of the four pump variables: flow, pressuredifference, speed and electromagnetic torque (apart from the possibilityof non uniqueness of solutions). Instead of torque, motor current orinput power can be used. Thus if viscosity is known, or if its influenceis sufficiently small, flow rate and pressure difference can beestimated from the motor speed and input power, which can be determinedfrom current and voltage measurements on the motor input leads.

The centrifugal blood pumps of previously described embodiments use ahydrodynamic bearing and can be constructed so that the variation withviscosity is sufficiently small to enable flow and pressure differenceestimation using signals derived from the coils 16, 17.

For this example a flow loop was set up consisting of the pump and 2.4 mof ⅜″ tubing giving a net fluid volume of 177 ml.

The fluid filled tubing was sunk into a water bath with a controlledheater. Temperature sensors were attached to the tubing to providevisual feedback on fluid temperature. Pressure taps were made on theinlet and outlet nozzles of the pump which interfaced to a differentialpressure transducer with digital display to measure pressure across thepump. A Clamp on Transonics flow probe and meter were used to measureflow rate and input power (motor supply voltage and current) wasmonitored via digital panel meters on the power supply. Pressure wasvaried by adjustment of a tubing clamp and motor speed by wuitableelectrical adjustment.

Two tests were conducted. The first with 5% saline, the second with redblood cell suspensions, haematocrit being 32%. In both cases thecirculating fluids were heated to 370C. 5% saline was chosen since itsviscosity is about that of water at 23deg C.

Flow rate, pressure head, pump speed and electrical input power weremeasured for both fluids.

Data for saline and blood was combined and correlated on a surface plotdescribing both flow rate as a function of motor speed and input poweras illustrated in FIG. 29.

Curve fitting of this plot produced the equationQ=20.29+4.731n(Pin)−55{square root}(n) where Q is flow rate in L/min,Pin is electrical input power to the motor in Watts and n is motor speedin rpm. The maximum error for this prediction was 4% for the combineddata. Pressure head across the pump was described by the relationshipΔP=−13.68−6.591n(Pin)+2.18e−5 (n)² with equivalent accuracy. Twodifferent rotor designs have been tested in this manner to date bothyielding similar accuracy curve fits of the form Q=a+b.ln(Pin)+c.{squareroot}(n) and of the form ΔP=a+b.ln(Pin)+c.(n)².

The viscosity of saline is approximately 1 mPas. The Viscosity of blood(Hct=32%) given pump shear rates of greater than 100 s⁻¹ is near 3 mPas.Blood viscosity varies from approximately 2.4 to 4.5 mPas over thephysiological range in question for shear rates greater than 100 s⁻¹.The variation in viscosity from 1 to 3 mPas produced a maximum error of4% in the prediction of flow rate.

The pump of FIGS. 1-6 has characteristics such that the model for flowrate prediction based on motor input power and speed is not greatlyaffected by variation in viscosity. This suggests for this pump it ispossible to determine flow with acceptable accuracy without using aseparate flow sensor.

The reasons for low error in prediction given change in viscosity arepostulated as follows: Firstly that the “flat” H-Q curves for this pumpgive small variation in pressure head for given flow rates. Secondly thenature of the hydro-dynamic hearing. Although the pump has relativelyhigh disc friction forces, which tend to be most sensitive to viscositychanges, the rotor in this case conserves energy by repositioning infree space according to the fluid viscosity. Thirdly, the size, wheresurface roughness is relatively smaller than for smaller higher speedpumps. Fourthly, allowing speed to vary around a set point due tochoosing a comparatively long time constant.

FIG. 30 illustrates the pump assembly 11 in cross section as utilisedwith example 1.

FIG. 31 illustrates in cross section the coil and magnet arrangementused in conjunction with example 1.

With reference to FIGS. 20, 30 and 31 iron yokes are placed outside thecoils to increase the magnetic flux and hence increase motor efficiency,and also to reduce stray magnetic fields in the body. The yokes arepositioned so that the axial magnetic force on the impeller is zero whenit is central in the housing cavity. Furthermore, the yokes are placedat considerable distances from the impeller to keep the negativemagnetic stiffness sufficiently low that is places only a smalladditional demand on the hydrodynamic suspension when the impellershifts away from the cavity mid-position.

Given the large distance to the yokes, a slotless winding andaxisymmetric yokes were chosen. The use of axisymmetric yokes implieszero “cogging” torque. The winding topology coil chosen is of “secondharmonic” type with just three coils, one per phase, in each of the bodyand cover windings. FIG. 31 depicts the cover winding. The body windingsalign axially with the cover windings but must be bent in severaldirections to avoid the volute and inlet. This second harmonic topologyavoids coil overlaps and is consequently neat and compact and gives lowcopper mass. However, it is less efficient than other winding optionswith greater coil mass.

The efficiency is increased by tilting the magnet alignment to an angleof 22.5° from the pump axis (as indicated in FIG. 30 by the magnethatching), intermediate between the 45° conical body and the flat cover.The cover coil and axial flux form an axial flux motor, and the bodycoil and flux are intermediate between an axial and radial flux motor.

The motor can be driven by a six-step, sensorless commutation inverter.Superimposed over the coils in FIG. 31 are magnets at an instant whenthe currents are switched from phases a and c conducting to phases b andc conducting (or v.v.). Parallel coil connection of the cover and bodycoils (each connected in star configuration) enables some redundancy, inthat the motor still runs with the loss of a coil.

The materials used were Ti-6A1-4V for the housing and impeller shell,high remanence NdFeB magnets (VACODYM 510 HR) embedded in the impeller,iron for the yokes (mild steel in prototypes but to be laminated siliconsteel) and varnished copper wire for the coils.

The measured negative magnetic stiffness of the teardrop impeller is−4000 N/m (±10%). The axial clearance gaps are 0.1 mm when the impelleris central (this is to match a 0.05 mm taper on the blades for thrustgeneration so that after a shift of 0.05 mm, the thrust forces aremaximal from one impeller face and negligibly small from the otherface). Thus if the impeller is shifted axially by the full amountpossible (as at start-up if axis vertical), then the magnetic force onthe impeller is 0.4 N force. This is less than the impeller weight of 46gforce, and is considered acceptable. If the yokes were any closer, theforce would be higher, increasing the risk of touchdown. Similarly, ifthe clearance gaps are increased to slacken manufacturing tolerances,then the maximal magnetic force can be increased.

The measured motor efficiency is between 45% and 48% curves, for speedsbetween 2000 rpm and 2500 rpm and motor output power between 3 and 7W.For example, at 2250 rpm and 3 W motor output (roughly ratedconditions), the copper loss was 1.7W, the eddy loss in the titanium was1.0 W. and the iron loss in mild steel yokes was 0.7 W, giving a motorefficiency of 47%.

With reference to FIG. 32 the example 1 can be applied to the preferredembodiment of FIG. 7 to 15 comprising pump assembly 200 incorporating anestimating and control system of the type described with reference toFIGS. 28 to 31.

With particular reference initially to FIG. 7 the pump assembly 200comprises a housing body 201 adapted for bolted connection to a housingcover 202 and so as to define a centrifugal pump cavity 203 therewithin.

The cavity 203 houses an impeller 204 adapted to receive magnets 205within cavities 206 defined within blades 207. As for the firstembodiment the blades 207 are supported from a support 208.

Exterior to the cavity 203 but forming part of the pump assembly 200there is located a body winding 209 symmetrically mounted around inlet210 and housed between the housing body 201 and a body yoke 211.

Also forming part of the pump assembly 200 and also mounted external topump cavity 203 is cover winding 212 located within winding cavity 213which, in turn, is located within housing cover 202 and closed by coveryoke 214.

The windings 212 and 209 are supplied from the electronic controller ofFIG. 32. Otherwise the structure is as described with reference to thethird embodiment.

Further Embodiments

In the forms thus far described top surfaces of the blades 8, 207 areangled at approximately 45° with respect to the longitudinal axis of theimpeller 100, 204 and arranged for rotation with respect to the internalwalls of a similarly angled conical pump housing. The top surfaces aredeformed so as to create the necessary restriction in the gap betweenthe top surfaces of the blades and the internal walls of the conicalpump housing thereby to generate a thrust which can be resolved to bothradial and axial components.

In the examples thus far the bottom faces of the blades 207 comprisesurfaces substantially lying in a plane at right angles to the axis ofrotation of the impeller and with their deformities define a gap withrespect to a lower inside face of the PUMP housing against which asubstantially only axial thrust is generated.

Other arrangements are possible which will also, relying on theseprinciples, provide the necessary balanced radial and axial forces. Sucharrangements can include a double support arrangement where the conicaltop surface of the blades is mirrored in a corresponding bottom conicalsurface. The only concern with this arrangement is the increased depthof pump which can be a problem for in vivo applications where sizeminimisation is an important criteria.

SUMMARY OF OPERATION PRINCIPLES

The estimation and control system described with reference to Example 1and the previous embodiments is “sensorless” in that it derives anestimate of relevant pump parameters from signals available from one ormore of the drive coils of the motor. Hence no separate sensor device isrequired to control the pump assembly in use.

It is hypothesized that the ability to control the pump assembly in thismanner to a sufficiently good approximation derives from shaping theimpeller of the pump so that a relatively flat head versus flowcharacteristic is obtained over the flow rate range expected and/orrequired of the pump, in use.

It is postulated that relative radial off-flow and lack of constraint ofthe fluid within the impeller derived from the relatively low number ofimpeller blades aids in achieving the relatively flat pumpcharacteristic curves as shown for example in FIGS. 3 and 13.

It is also postulated that, in the embodiments described in thespecification, the impeller blades are arranged to guide fluid carefullythrough the rotor so as to reduce re-circulation. There are alsorelatively large gaps between the blades so that the fluid is relativelypoorly constrained leading to loosely constrained flow of fluid withinthe pump housing.

EXAMPLE 2

With reference to FIGS. 33 to 39 a specific example of a particularlypreferred rotor, centrifugal flow pump assembly incorporating the rotorand a control system therefor will now be provided.

The rotor 500 of this example is illustrated in FIG. 35 and is arrangedto operate within a housing structure as previously described in thisspecification with reference to FIG. 7. The rotor 500 is urged to rotateby an electromagnetic field supplied via coil structures again aspreviously described with reference to FIG. 7. The control system whichmaintains control over the operation of the rotor within the housing isdetermined by a non-contact estimation and control system as previouslydescribed in this specification but further subject to an optimalpumping condition strategy as will be described below.

With particular reference to FIG. 35 it will be noted that the impellerblades 501 are held in mechanical relationship with each other by struts502.

By increasing the smallest radius from the centreline to the blades(i.e. to the nose of the blades) at the top and not at the bottom of theimpeller, an axial thrust force can be imposed on the impeller towardthe bottom. This arrangement can be carefully designed so as to bias theload to the bottom bearing and relieve the top bearing which is morehighly loaded (in that it must resist both axial and radial loads).

Operation Region for the Pump

With reference to FIG. 34 the pump of this example is arranged to followan HQ curve that does not roll-off towards shut-off. That is, if thepressure head (H) developed by the pump at any given operating speed (N)is plotted against flow rate delivered (Q), then at low flow rates (andeven at zero flow) there is no loss of head compared with the headdeveloped at the nominal operating point. Typically in other centrifugalpumps of the prior art pressure head developed increases with increasingflow rate from zero or “shut-off” to a point of inflection in the HQcurve then head reduces with further increases in flow rate. It isnormal practice in the prior art to operate a pump to the right of theinflection point to avoid instabilities known as surge which occurbecause at a given pump speed a pressure head required might be met byone of two flow rates delivered by the pump—one at either side of theinflection point. To the right of the inflection point, typically the HQcurve falls steeply.

In the pump of this example, since there is no inflection point in theHQ curve, the pump can be operated stably throughout its entire range offlow rates. This means that the pump is operating in the flattest partof the HQ curve and enables better prediction of flow and pressure fromparameters which may be attained readily from motor performancecharacteristics (viz.: Voltage, current and speed).

Factors which contribute to the flat HQ curve of the pump of thisexample, with an absence of an inflection point, include near-radialoff-flow from the impeller, low specific-speed design of the pump and alow number of impeller blades.

An optimal control strategy will now be described with reference toFIGS. 33 to 39.

Optimal Control Strategy

It is the aim of the rotary blood pump and its associated control systemof Example 2 to restore normal cardiac output levels such that thedemand for perfusion is supplied by pumping as much blood from theventricle as is returned to it from the lungs.

Rate responsive control of the pump is described in this example todetermine the optimum point for unloading the heart while at the sametime avoiding over pumping leading to suction or under pumping leadingto regurgitation during the varying physiological climate of every daylife.

Since the pump has no valves, there is a possibility of back flow whenthe pump speed is low. FIG. 36 shows the normal direction for flow ofblood through the pump from the left ventricle to descending aorta. Thispoint changes with preload (left ventricle pressure) and afterload(arterial pressure) across the pump. Furthermore, as pump speed isincreased the aortic valve will eventually remain closed and additionalincreases in speed will cause collapse of the ventricle.

The rotary blood pump is sensitive to pre load and after load if thepump speed set point has no feed back. Instantaneously increasing thepressure head across the pump will cause an increase in impeller speedand decrease in electrical input power and pump flow rate. Decreasing itwill have opposite effects.

In this example the time constant of the control system is set to berelatively slow to the extent that disturbances induced in the speed ofblood flow by the action of the heart will be counteracted by thecontrol system relatively slowly. The resulting variation in speed ofthe impeller, in use, is then utilised to calculate an estimate of theoperating point to an improved level of accuracy.

The long time constant means that instantaneous pump speed andelectrical input power will vary cyclically under the influence of thepumping action of the heart or, in other words, will be modulated by theheart beat.

In this example the time constant of the control system is set to begreater than the rotational inertial time constant of the impeller.Specifically, in this example, the time constant is set at 5 secondswhich is longer than one cardiac cycle.

Optimal Pumping and Avoiding Over Pumping

If pump speed is set such that maximal unloading of the ventricle isachieved and venous return is reduced as in the case from exercise toresting, over pumping from the ventricle will result in suction andcollapse of the ventricle may occur.

As the pump speed is increased the ventricle empties and the pressure inthe ventricle during systole decreases. This is shown in FIG. 37 as dPis reduced from dp1 to dP3 during systole. If the amount of emptying ofblood from the ventricle matches the amount of filling, the pump is onthe verge of producing suction or negative pressures in the ventricle.Beyond the point that the aortic valve remains closed and the peak leftventricular pressure during systole continues to decrease and suctionwill begin to occur during the diastolic phase.

Further increases in pump speed will cause the peak left ventricularpressure to become so low that the ventricle walls will occlude bloodflow through the inlet cannula over the entire cardiac cycle, evenduring systole. Suction should be avoided even during diastole. Theoptimum point of pumping is just allowing the aortic valve to open. Overpumping is considered increasing pump flow beyond this point.

The solution to detection of the point of optimal pumping in thisexample lies in the time domain.

The point at which the aortic valve just remains closed is the point oftotal assist given the name OCA (optimal cardiac assistance) This is thepoint at which minimum head pressure across the pump begins to rise withincreasing pump speed. In other words during systole the left ventriclepeak pressure begins to decrease as average pump speed is increased.

Therefore for a given preload, afterload and contractile strength of theventricle there will be a point where optimum unloading of the ventricleoccurs. Increase in pump speed beyond this point will result in collapseof the ventricle. This minimum pressure across the pump during systolewill produce a maximum flow through the pump, maximum torque on theimpeller and minimum instantaneous speed.

Therefore pumping at the point of optimal cardiac assistance andavoiding over pumping, the control algorithm should maintain minimumpump speed such that the minimum head pressure across the pump does notincrease. Therefore the new desired set point Nnew to hold the optimalcardiac assistance point can he defined by the old speed value Noldreduced by a factor proportional to the increase in minimum systolichead pressure (ΔHsys) beyond the minimum possible head pressure (ΔHmin)Kp is the proportional constant. This is described by equation 1.Nnew=Nold−[Kp*(ΔHsys−ΔHmin)]  equation 1

The instantaneous head pressure can be estimated by non-contact methodsas previously described in this specification with reference to Example1.

This is a simple control equation that can be readily implemented in anembedded microcontroller system.

Avoiding Under Pumping

The other boundary condition of under pumping occurs when flows throughthe pump become negative with diastole. Regurgitation can causestagnation of blood and lead to thrombus formation as well as increasingatrial pressures leading to pulmonary adaema.

Regurgitant or negative flow in the pump begins to occur as pump setpoint speed is decreased to the extent where levels and phase lagsbetween pump outlet and inlet pressures during diastole cause flowreversal.Nregurg=N(t) for Qdiastole=0L/min  equation 2where N(t) is the instantaneous impeller rotational speed, Qdiastole isthe minimum flow rate through the pump during diastole and Nregurg isthe minimum speed at which non regurgitant flow occurs. Flow rate can beestimated by non-contact methods as previously described as a functionof motor speed and input power as shown earlier with reference toExample 1.SUMMARY OF OPERATIONAL FORCES EXPERIENCED

In practical operation of Example 2 the rotor 500 should be made tooperate such that blood flow is adequate in accordance with theconstraints and the optimal control strategy described above. Inaddition, whilst in operation, the rotor 500 ideally should never makecontact with the inside walls of the housing in which it rotates. Shouldsuch contact be made then the control system should be able to recoverfrom this condition so as to return the rotor to an operationalcondition and, in addition, damage sustained during a touchdown must beminimised so that, upon return to normal operation after touchdown,there is no effect on steady state operation.

Touchdown is countered by ensuring that there is sufficient restoringhydrodynamic force exerted upon the rotor 500 to counteract anydisturbing force experienced by the rotor 500 such that probability oftouchdown is reduced to a sufficiently low value.

Broadly, for the centrifugal pump structure described with reference tothis example and having a rotor 500 of the type described with referenceto this example it has been found that worst case restoring forces occurwhen the rotor is rotating in a low viscosity medium and running at itslowest speed. For example, running at approximately 1800 rpm in a bloodsubstitute representing the lowest viscosity likely to be encountered inpractice the axial restoring force available is approximately 2 Newton.The corresponding radial restoring force is approximately 0.5 Newtonunder these conditions.

In a more usual and expected operating environment with blood viscosityof the order of 2.5 mPs and with the rotor running at approximately 3000rpm the axial restoring force available is approximately 9 Newton whilstthe radial restoring force is approximately 2.25 Newton. If the speed isreduced to around 2400 rpm then the axial restoring force isapproximately 5.3 Newton and the radial restoring force is approximately1.3 Newton.

An expected typical steady state disturbing force can be of the order of0.45 Newton including the effects of gravity upon rotor 500. Magneticfield disturbances arising from the drive mechanism can add a further0.1 Newton of disturbing forces. Allowing for acceleration effects onthe entire assembly in vivo and in use it is expected that typicalmaximum disturbing forces encountered by rotor 500 will be of the orderof 1 Newton.

Use of a shrouded rotor as described earlier in the specification candouble the radial resistance force available to counteract disturbingforces.

In addition appropriate coatings and/or structure materials will be usedon the respective rotor 500 and at least inner walls of the housing tominimise damage and/or damaging effects arising from a touchdown.

Coatings and/or inherent structural materials will be selected so as toreduce the friction co-efficient between the rotor and the casing andalso to reduce specific damage such as gouging.

Current particularly preferred materials for this purpose includeamorphous carbon based materials or microcrystalline carbon basedmaterials and titanium nitride. In a particular form it has been foundadvantageous to have different materials on opposing surfaces such as,for example, titanium nitride on one of the surfaces and the carbonbased materials on the other.

Carbon based material against carbon based material correspondingsurfaces has been found to give a very low co-efficient of friction(typically 0.05) and low damage. Conversely, titanium against titaniumhas been found to give the reverse effect and is not recommended.

Further particular preferred coating arrangements are as follows.

Application of coatings to the Ti-6A1-4V substrate of a blood pump.

-   -   These coatings specifically include carbon with a graphitic        microcrystalline structure, prepared using unbalanced magnetron        sputter deposition and amorphous carbon coatings.    -   These coatings provide a biocompatible lining for the device,        with enhanced hardness, high elastic recovery under impact        conditions, low friction coefficients of <0.06 under lubricated        conditions, and high wear resistance.    -   Also coatings applied using plasma immersion ion implantion of        nitrogen, titanium nitride and carbon, or combinations of these        treatments, for enhancement of hardness, improved elastic        recovery under impact conditions, low friction and high wear        resistance.

Other potential candidates include pyrolitic carbon, illumina, zirconiaor combinations thereof.

Overall the desirable characteristics to be achieved by the coatingsare.

-   -   1. Zero eddy current loss in the casing;    -   2. A hard material on at least one surface to give a good        bearing property; and    -   3. The materials must be biocompatible, particularly in relation        to blood contact.

More generally in relation to modification to surfaces attention can bepaid to other characteristics of the surface structure of both the rotor500 and the at least inner wall of the housing with a view to providingflexibility or other dynamic characteristics which can aid hydrodynamicbearing behaviour. Elasto-hydrodynamic bearings are one such arrangementwhich can be achieved by further attention to the materials of which therotor and/or housing inner walls are constructed as will now bedescribed.

Elastohydrodynamic Bearings (EHD Bearings)

Elastohydrodynamic (EHD) bearings rely on the principle that if forcesapplied to bearings are sufficiently large then the bearing surface willdistort. This distortion can lead to a greater efficacy of the bearingsallowing greater loads to be carried for a given bearing dimension ofcourse the magnitude of force necessary to distort the bearing surfacemust be viewed relative to the modulus of elasticity of the materialfrom which the bearing surface is manufactured. For EHD to be applicableto rotary blood pumps, the modulus of elasticity of the bearing surfacesmust be low in order for them to be distorted by forces of a few Newtonsmagnitude. For these reasons polymeric materials such as polyurethane orsilicone may prove acceptable materials from which to fabricate thebearing surfaces (these materials should sit on a harder substrate of,say, titanium or a ceramic). The fundamental shape of the hydrodynamicbearing, including particularly the “deformed surfaces” would need toremain substantially the same as for the hydrodynamic bearingspreviously described in this application for use in a blood pump.

BIO-EHD Bearings

An alternative approach is to allow tissue overgrowing the bearingsubstrate to act as the EHD component. The substrate may be porous suchthat it allows a pseudoneointimal cell lining to grow into the pores onthe substrate surface. It is commonly reported that the pseudoneointimallining thickness is stable and around one cell deep. The advantage ofusing a “bio-EHD” component on the surface is that damage to the EHDcomponent may regenerate within a few hours of damage occurring.Possible drawbacks may be a tendency for a bio-EHD component to sustaindamage under relatively low shear stresses commonly seen in a rotaryblood pump and for sections of bio-EHD to be stripped away formingpotentially dangerous embolii. This may be countered by additionalsurface treatments which promote stability of the pseudoneointima. Onceagain, the fundamental shape of the bearing, that is the “deformedsurfaces” should remain substantially the same as for the hydrodynamicbearings of the blood pump embodiments previously described.

The above describes only some embodiments and some examples of a rotaryblood pump and control system therefor and modifications, obvious tothose skilled in the art, can be made thereto without departing from thescope and spirit of the present invention.

INDUSTRIAL APPLICABILITY

The pump assembly 1, 200 is applicable to pump fluids such as blood on acontinuous basis. With its inherently simple mechanical and controlstructure it is particularly applicable as an in vivo heart assist pump.

The pump assembly can also be used with advantage for the pumping ofother fluids where damage to the fluid due to high shear stresses shouldbe avoided or where leakage of the fluid should be prevented with a veryhigh degree of reliability—for example where the fluid is a dangerousfluid.

1. A method of hydrodynamically suspending and controlling an impellerwithin a rotary pump for support in at least one of a radial and axialdirection; the method comprising: incorporating a deformed surface in atleast part of the impeller so that, in use, a thrust is created betweenthe deformed surface and an adjacent pump housing during relativemovement therebetween; and maintaining the speed of rotation of theimpeller within a range whereby the impeller, in use, substantiallyresists five degrees of freedom of movement with respect to the pumphousing without any external intervention.
 2. The method of claim 1wherein the deformed surface includes a taper.
 3. The method of claim 2wherein the taper is arranged so that there is a larger gap at a leadingedge thereof between the impeller and the pump housing than at atrailing edge thereof.
 4. An estimation and control system for a pump;the pump of the type having an impeller located within a pump cavity ina pump housing; the housing having a fluid inlet in fluid communicationwith the cavity; the housing having a fluid outlet in fluidcommunication with the pump cavity; the impeller urged to rotate aboutan impeller axis so as to cause fluid to be urged from the inlet throughthe pump cavity to the pump outlet; the impeller urged to rotate by animpeller drive; the impeller supported for rotational movement by animpeller support; the pump maintained at or near a predeterminedoperating point by a controller acting on the impeller drive; thecontroller receiving as input variables at least a first input variablederived from the impeller drive; the controller receiving at least asecond input variable also derived from the impeller drive; thecontroller thereby calculating an estimate of the operating point to anapproximation of predetermined accuracy relying on signals availablefrom the impeller drive; the controller controlling the pump bycomparing the predetermined operating point with the estimate of theoperating point; and wherein the pump is arranged to operate accordingto a relatively flat pressure versus flow rate characteristic.
 5. Theestimation and control system of claim 4 wherein there is no inflexionpoint in the pressure versus flow rate characteristic at or near thepredetermined operating point.
 6. The estimation and control system ofclaim 4 wherein the pump includes near-radial off-flow from theimpeller.
 7. The estimation and control system of claim 4 wherein thepump has a low specific speed.